JP5195491B2 - Control device for internal combustion engine - Google Patents

Description

  The present invention relates to a control device for an internal combustion engine capable of directly supplying a plurality of fuels having different ignitability to a fuel chamber.

  For example, Patent Document 1 uses two types of fuels with different ignitability typified by octane number (RON) as a technique for suppressing combustion noise during high-load operation, which is a problem of the premixed compression self-ignition method. It has two in-cylinder direct injection injectors per cylinder, and injects low ignitable fuel and high ignitable fuel so that they do not overlap in the combustion chamber, and provides a low ignitable fuel region and a high ignitable fuel region in the combustion chamber. The forming technique is known. According to this, it becomes possible to suppress steep combustion by forming a phase difference between two gas mixture regions having different ignitability at the ignition timing in one cycle.

JP-A-2005-139945

  However, in this Patent Document 1, there is a problem that, at a load where the injection amount of low ignitable fuel is small, the injection amount is too small, the air-fuel mixture becomes excessively diluted, and the thermal efficiency deteriorates due to the reduction in combustion efficiency. .

  Therefore, the control apparatus for an internal combustion engine according to the present invention includes a first operation mode in which low ignitable fuel and high ignitable fuel are mixed in the combustion chamber, and different spaces in the combustion chamber in which the low ignitable fuel and high ignitable fuel are mixed. And switching between the first operation mode and the second operation mode at the boundary of the load when the amount of low ignitable fuel exceeds a predetermined value. It is a feature.

  In the second operation mode, there is a fuel amount necessary for the low ignitable fuel to form a relatively homogeneous air-fuel mixture with λ (excess air ratio) = 2 to 3, and the load at which the fuel amount is obtained. In the following, when the second operation mode is performed, the air-fuel mixture becomes excessively diluted (λ> 3), and the unburned ratio increases, so that the thermal efficiency is deteriorated.

  According to the present invention, the optimal switching time between the first operation mode and the second operation mode is set, and deterioration of thermal efficiency can be prevented.

BRIEF DESCRIPTION OF THE DRAWINGS Explanatory drawing which showed an example of the system structure of the control apparatus of the internal combustion engine which concerns on this invention. Explanatory drawing which showed typically an example of the variable compression ratio mechanism applied to this invention. Explanatory drawing which showed typically the mode of the 1st combustion mode implemented by the control apparatus of the internal combustion engine which concerns on this invention. Explanatory drawing which showed typically the mode of the 2nd combustion mode implemented by the control apparatus of the internal combustion engine which concerns on this invention. Explanatory drawing which showed the state of the top dead center position in 2nd combustion mode typically. Explanatory drawing which showed typically an example of switching control of the 1st combustion mode and 2nd combustion mode implemented by the control apparatus of the internal combustion engine which concerns on this invention. The characteristic view which shows the correlation with the octane number (RON) of a low ignitability fuel, and 2nd combustion mode execution load. The characteristic view which shows the correlation with the tank remaining amount of low ignitability fuel, and 2nd combustion mode execution load. The timing chart which showed the example of control when there is enough quantity of reforming fuel (low ignition quality fuel). The timing chart which showed the example of control when the quantity of reformed fuel (low ignition quality fuel) is insufficient. The timing chart which showed the example of control when the octane number (RON) of reformed fuel (low ignitability fuel) is high enough. The timing chart which showed the example of control when the octane number (RON) of reformed fuel (low ignition property fuel) is low. 6 is a timing chart showing an example of control when the intake pressure is constant and the amount of reformed fuel (low ignitable fuel) is sufficient. 6 is a timing chart showing an example of control when the amount of reformed fuel (low ignitable fuel) is insufficient with a constant intake pressure.

  Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings. FIG. 1 is a system configuration diagram of a control device for an internal combustion engine according to the present invention.

  As shown in FIG. 1, the internal combustion engine 1 includes a first common rail fuel injection system including a first common rail 2, a first fuel injection valve 3, and a first fuel pump 4, a second common rail 5, and a second common rail. And a second common rail fuel injection system having the fuel injection valve 6 and the second fuel pump 7 as components. The first fuel injection valve 3 and the second fuel injection valve 6 directly inject fuel into the combustion chamber 8, and the fuel injection pressure is changed by changing the set fuel pressure in each of the common rails 2 and 5, respectively. Variable control is now possible. The piston 9 that defines the combustion chamber 8 has a cavity 11 formed at the center of the piston crown surface 10. In addition, 12 in FIG. 1 is a spark plug.

  The fuel in the fuel tank 15 is supplied to the first common rail fuel injection system by the first fuel pump 4. On the other hand, the fuel in the reformed fuel tank 16 is supplied to the second common rail fuel injection system by the second fuel pump 7. An ignitable fuel different from the fuel in the fuel tank 15 is stored in the reformed fuel tank 16. The fuel in the reformed fuel tank 16 is obtained by reforming the fuel in the fuel tank 15 into an ignitable fuel different from the fuel in the fuel tank 15 by a fuel reformer 40 (described later).

  In the present embodiment, the fuel in the fuel tank 15 is a commercially available light oil, and the fuel in the reformed fuel tank 16 is reformed in fuel properties to a higher octane number than the light oil, and has a lower ignitability than the light oil. Of fuel is stored. That is, in the present embodiment, from the first fuel injection valve 3. Highly ignitable (low octane number) fuel is injected from the second fuel injection valve 6. Low ignitability (high octane number) fuel is injected.

  The internal combustion engine 1 has a supercharger 20 including a turbine 21 installed in the exhaust passage 18 and driven by exhaust gas, and a compressor 22 installed in the intake passage 17 and driven by the turbine 21. The turbocharger 20 of the present embodiment adopts a variable displacement type, and in the low speed range, the variable nozzle (vane) 23 provided on the turbine 21 side is narrowed to increase the turbine efficiency, and in the high speed range. By opening the variable nozzle 23 and increasing the turbine capacity, a high supercharging effect is obtained in a wide operation region.

  In the intake passage 17, an air flow meter 25 and an air cleaner 26 are disposed on the upstream side of the compressor 22 of the supercharger 20, and an intercooler 27, a throttle valve 28, and a swirl control valve 29 are interposed on the downstream side.

  The exhaust passage 18 is provided with an EGR passage 31 that branches from between the internal combustion engine 1 and the turbine 21 and connects to the intake passage 17. An EGR valve 32 and an EGR cooler 33 are interposed in the EGR passage 31. The EGR valve 32 is, for example, an electronically controlled type using a step motor, and controls the amount of exhaust gas recirculated to the intake side according to the opening, that is, the EGR amount sucked into the internal combustion engine 1 It is. The EGR cooler 33 cools the exhaust gas recirculated from the exhaust passage.

  In the exhaust passage 18, an oxidation catalyst 35, a NOx trap catalyst 36, and a DPF (diesel particulate filter) 37 are sequentially provided on the downstream side of the turbine 21. The oxidation catalyst 35 oxidizes HC and CO in the active state. The NOx trap catalyst 36 traps NOx (nitrogen oxide) in exhaust during lean fuel, and reduces and purifies NOx trapped during stoichiometric combustion or rich combustion using HC and CO in exhaust as reducing agents. Is. The DPF 37 collects PM (particulate matter) in the exhaust gas, and the collected PM is combusted by regeneration control for increasing the exhaust gas temperature.

  The above-described fuel reformer 40 is a fuel converter that uses a catalyst (platinum catalyst or the like) and the exhaust heat of the internal combustion engine 1. A low octane fuel represented by n-paraffin is subjected to a dehydrogenation reaction to degenerate the engine. Convert to high octane fuel suitable for high load operation. The fuel in the fuel tank 15 is supplied to the fuel reformer 40 by a third fuel pump 41 and a fuel supply valve 42. The combustion reformer 40 is provided with a combustion reformer temperature sensor 43 for detecting the temperature (reforming temperature). The reformed gas generated by the fuel reformer 40 is stored in the reformed fuel tank 16 via the gas compressor 44 and the aggregator 45. The amount of reformed fuel stored in the reformed fuel tank 16 is detected by a fuel sensor 46.

  As sensors for detecting various states, the air flow meter 25 for detecting the intake air amount, a rotational speed sensor 51 for detecting the engine rotational speed, an accelerator opening sensor 52 for detecting the accelerator opening, and a water temperature sensor 53 for detecting the cooling water temperature. An intake pressure sensor 54 for detecting the intake pressure downstream of the throttle valve 28; an intake air temperature sensor 55 for detecting the intake air temperature downstream of the throttle valve 28; an air / fuel ratio sensor 56 for detecting the exhaust air / fuel ratio at the inlet of the oxidation catalyst 35; An air-fuel ratio sensor 57 for detecting the exhaust air-fuel ratio at the outlet of the trap catalyst 36, a temperature sensor 58 for detecting the exhaust temperature of the DPF 37 inlet, a differential pressure sensor 59 for detecting the pressure difference between the inlet and outlet of the DPF 37, and the exhaust temperature of the DPF 37 outlet A temperature sensor 60 for detecting fuel, the combustion reformer temperature sensor 43, and the fuel sensor 46. Is provided, the detection signals from these sensors and the like are input to the C / U (control unit) 65.

  The C / U 65 is composed of a microcomputer comprising a CPU and its peripheral devices. The fuel injection amount and the injection timing are set based on the detection signals from the various sensors described above, and the fuel injection valves 3, 6 are set. While controlling the drive, the opening control of the throttle valve 28, the EGR valve 32 and the swirl control valve 29 is performed.

  Moreover, the internal combustion engine 1 in this embodiment is a variable compression ratio that can change the compression ratio to an arbitrary compression ratio by changing the combustion chamber volume at the top dead center using a multi-link piston-crank mechanism. It has a mechanism.

  As shown in FIG. 2, the variable compression ratio mechanism includes an upper link 75 having one end connected to a piston 9 that slides in a cylinder 72 of a cylinder block 71 via a piston pin 74, and the other end of the upper link 75. And a lower link 79 connected to the crankpin 78 of the crankshaft 77 in a rotatable manner, and further connected to the lower link 79 to limit the degree of freedom of the lower link 79. A control link 81 having one end connected via a pin 80 and the other end supported by the internal combustion engine body so as to be swingable. The swing support position of the control link 81 is the eccentricity of the control shaft 82. The cam portion 83 is variably controlled.

  The control shaft 72 is disposed in parallel with the crankshaft 77 and is rotatably supported by the cylinder block 71. The control shaft 72 is driven in the rotational direction by an actuator 85 made of an electric motor via a gear mechanism 84, and its rotational position is controlled.

  In the variable compression ratio mechanism configured as described above, the swing support position of the lower end of the control link 11 changes depending on the rotational position of the control shaft 12, that is, the position of the eccentric cam portion 13, and the initial posture of the lower link 9 changes. Along with this, the top dead center position of the piston 9, and the compression ratio changes accordingly.

  The internal combustion engine 1 uses a first combustion mode (first operation mode) as shown in FIG. 3 and a second combustion mode (second operation mode) as shown in FIG. They are switched accordingly. In the first combustion mode, the highly ignitable fuel injected from the first fuel injection valve 3 and the low ignitable fuel injected from the second fuel injection valve 6 are mixed in the combustion chamber 8. In the second combustion mode, the low ignitable fuel and the high ignitable fuel are divided and distributed in different spaces in the combustion chamber 8, and the high ignitability is obtained in the region above the cavity 11 (low RON region) of the piston 9. Fuel (low RON fuel) is distributed, and low ignitable fuel (high RON fuel) is distributed in a region outside the cavity 11 (high RON region) in the combustion chamber 8. Note that spark ignition means (reference numeral 12 in FIG. 1) may be provided in the combustion chamber, and combustion may be performed by spark ignition in a load range higher than the load range in which the second combustion mode is performed.

  The first combustion mode will be described. In the first combustion mode, if the self-ignition timing is too early, the combustion becomes steep and the sound vibration deteriorates. Further, if the ignition timing is advanced from a predetermined ATDC, not only the sound vibration but also the efficiency is deteriorated (cooling loss is deteriorated). On the other hand, if the self-ignition timing is too late, the efficiency deteriorates (exhaust loss increases). Therefore, in order to obtain heat generation characteristics (not steep) and timing (not too late but not too early) that achieve both sound vibration and efficiency, a lot of highly ignitable fuel (low RON fuel) is used on the low load side. Increase low ignitable fuel (high RON) on the load side.

  Further, as the load increases, the fuel increases and λ (excess air ratio) shifts to the rich side (the excess air ratio value decreases). When λ becomes rich, the self-ignition timing is advanced and the sound vibration is worsened (if diesel (diffusion combustion), the injection timing is retarded, but compression self-ignition cannot do so). Although it is conceivable to retard the ignition timing by increasing the octane number (RON), if the ignition timing is retarded too much, the efficiency deteriorates in the first place (exhaust loss increases). It will hit the limit that can not continue compression self-ignition with the load of.

  Next, the second combustion mode will be described. In the second combustion mode, since the ignition timing is shifted for each air-fuel mixture formed with fuel having different octane numbers (RON), it is possible to prevent the combustion from becoming steep. In the second combustion mode, when the low ignitable fuel (high RON fuel) decreases on the low load side, the air-fuel mixture (high RON mixture) due to the low ignitable fuel becomes too dilute and combustion is difficult to occur. That is, in the second combustion mode, there is a fuel amount necessary for the low ignitable fuel to be able to form a relatively homogeneous mixture with λ = 2 to 3, and the fuel amount is less than the load at which the fuel amount is reached. When the two-combustion mode is implemented, the air-fuel mixture becomes excessively diluted (λ> 3), and the unburned ratio increases, so that the thermal efficiency is deteriorated. Then, the intended combustion (non-steep heat generation) cannot be obtained, and the unburned fuel is discharged as it is. Moreover, even if the load for switching to the second combustion mode is too high, it becomes difficult to suppress the combustion noise. Therefore, there is a limit of establishment of the second combustion mode on the low load side.

  Therefore, in the present embodiment, the second combustion mode is stopped at a load at which the air-fuel mixture (high RON air-fuel mixture) of the low ignitable fuel outside the cavity 11 is larger than a predetermined λ (excess air ratio). . That is, as shown in FIGS. 5 and 6, in the first embodiment, the second combustion mode is performed when the condition X · Vs> N is satisfied. Here, X is a correction coefficient, Vs is the volume of the region outside the cavity 11 in the combustion chamber 8 at the top dead center, and N is a load for switching between the first combustion mode and the second combustion mode.

  In other words, the time when the air-fuel mixture can be formed with the low ignitable fuel alone is determined by the shape of the combustion chamber 8, and the time is set as the optimum switching time between the first combustion mode and the second combustion mode, thereby improving the thermal efficiency To prevent the deterioration.

  Further, since the injection amount of the low ignitable fuel (high RON fuel) is determined according to the load, the switching load N is determined by Vs and the injection amount of the low ignitable fuel (high RON fuel). be able to. That is, in the first embodiment, the first combustion mode and the second combustion mode are switched on the basis of the load when the amount of low ignitable fuel exceeds a predetermined value.

  Therefore, in such a 1st embodiment, while being able to prevent deterioration of thermal efficiency, suppression of combustion noise can be aimed at.

  Further, in the second combustion mode, the load that becomes the amount of fuel necessary for the low ignitable fuel to form a relatively homogeneous mixture with λ = 2 to 3 is determined by the shape of the combustion chamber 8. . Specifically, the larger the volume of the combustion chamber 8 at the top dead center, the larger the amount of fuel necessary for the low ignitable fuel to form a relatively homogeneous mixture with λ = 2 to 3. That is, the “predetermined value” in the “load when the amount of low ignitable fuel exceeds a predetermined value” increases as the volume of the combustion chamber 8 at the top dead center increases.

  In addition, when the compression ratio is changed by the variable compression ratio mechanism, the top dead center volume decreases as the compression ratio increases, so that a relatively homogeneous mixture with a low ignitable fuel of λ = 2 to 3 can be formed. The amount of fuel needed to become less. That is, the “predetermined value” in the “load when the amount of low ignitable fuel exceeds a predetermined value” decreases as the compression ratio increases. In other words, the load for switching between the first combustion mode and the second combustion mode increases as the compression ratio decreases.

  That is, the combustion noise is suppressed by lowering the compression ratio under the condition where the combustion noise is desired to be suppressed by the second combustion mode but the switching load between the first combustion mode and the second combustion mode is not reached.

  Further, when the low ignitable fuel and the high ignitable fuel are distributed in different spaces in the combustion chamber 8 as in the second combustion mode, a configuration in which the cavity 11 is provided in the piston crown surface 10 of the piston 9 is provided. It is advantageous. That is, by using the piston 9 having the cavity 11 and injecting different fuels into the inside and outside of the ridge line (edge) of the cavity 11, the fuel with a different octane number (RON) compared to a combustion chamber with a flat piston crown surface. It is easy to stratify the air-fuel mixture. That is, by providing the cavity 9 in the piston 9, it becomes easy to distribute the low ignitable fuel and the high ignitable fuel separately in different spaces in the combustion chamber 8.

  The “predetermined value” in the “load when the amount of low-adhesion fire fuel exceeds a predetermined value” is set so as to increase as the volume of the cavity 11 increases when the bore diameter is constant. The In other words, the “predetermined value” in the “load when the amount of low-ignition fire fuel exceeds a predetermined value” is the outside of the cavity 11 of the combustion chamber 8 at the top dead center when compared with a constant bore diameter. The volume is set so as to decrease as the volume of the battery increases.

  Hereinafter, other embodiments of the present invention will be described. In the load ranges shown in the following drawings, compression self-ignition combustion is performed over the entire range. The intake air amount is not basically adjusted, and the output is controlled by the fuel injection amount as in a diesel engine. Therefore, in the internal combustion engine 1, the fuel amount (fuel injection amount) increases as the load increases, and λ (excess air ratio) changes to the rich side.

  As shown in FIGS. 7 and 8, the load for switching between the first combustion mode and the second combustion mode is the octane number (RON) or fuel amount of the reformed fuel (low ignitable fuel) obtained by the fuel reformer 40. You may make it change according to.

  Since the fuel reforming performed in the fuel reformer 40 is an endothermic reaction (for example, dehydrogenation reaction), the fuel reformer 40 is configured to exchange heat with the exhaust gas (in terms of equilibrium, higher temperature and lower pressure are better). advantageous). For this reason, for example, since the exhaust gas temperature varies depending on the load, the octane number (RON) and the amount of fuel of the obtained reformed fuel vary depending on the load.

  Therefore, as shown in FIG. 7, the higher the octane number (RON) of the reformed fuel (low ignitable fuel) obtained by the fuel reformer 40, the smaller the execution load (switching load) in the second combustion mode. You may do it. Further, as shown in FIG. 8, the second combustion mode is executed as the remaining amount of the reformed fuel (low ignitable fuel) obtained by the fuel reformer 40 (the remaining amount of the reformed fuel tank 16) increases. The load (switching load) may be reduced.

  9 and 10, the combustion mode may be controlled according to the amount of reformed fuel (low ignitable fuel) generated by the fuel reformer 40. FIGS. 9 and 10 are timing charts showing examples of control when the amount of reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficient and when it is insufficient.

  When the amount of reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficient, as shown in FIG. 9, the first combustion mode and the second combustion mode are switched at a predetermined switching load. Perform switching.

  On the other hand, when the octane number (RON) in the reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficient but not sufficient, the control is performed as shown in FIG. . That is, in order to save low ignitable fuel (high RON fuel), the fuel injection amount of low ignitable fuel (high RON fuel) should not be increased so much even if the load increases (as a result in this embodiment). The average octane number hardly changes), and an increase in load corresponds to an increase in highly ignitable fuel (low RON fuel). On the other hand, the fact that the rapid combustion mitigation by the low ignitable fuel (high RON fuel) is not sufficiently performed is dealt with by stopping the supercharging. Note that this embodiment is based on the premise that supercharging is performed as the load increases in the first combustion mode. The decrease in output due to the suspension of supercharging corresponds to the increase in fuel amount. This is indicated by a stepwise increase in the injection period and total fuel quantity of high ignitable fuel (low RON fuel). When the effect of steep fuel mitigation by supercharging stop is crushed, the limit of homogeneous compression self-ignition is reached.

  11 and 12, the combustion mode may be controlled according to the octane number (RON) of the reformed fuel (low ignitable fuel) generated by the fuel reformer 40. 11 and 12 show examples of control when the octane number (RON) of the reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficiently high and when the octane number (RON) is low, respectively. It is a timing chart.

  When the octane number (RON) of the reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficiently high, as shown in FIG. 11, the first combustion mode and the second combustion mode are set at a predetermined switching load. Switch the combustion mode.

  On the other hand, when the octane number (RON) in the reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is low, control is performed as shown in FIG. That is, even if the low ignitable fuel (high RON fuel) is increased, the average RON cannot be sufficiently increased. Since it is not necessary to increase the average octane number as much as possible within the range where the effect of steep combustion mitigation by weakening the supercharging is obtained, the amount of low ignitable fuel (high RON fuel) as a result in this embodiment is as shown in FIG. Is also decreasing. The amount of fuel increases as the supercharging is weakened, and the efficiency deteriorates accordingly.

  FIGS. 13 and 14 show a case where the combustion mode is controlled in accordance with the amount of reformed fuel (low ignitable fuel) generated by the fuel reformer 40 and adjusted by exhaust gas recirculation instead of supercharging. It is a timing chart. That is, FIGS. 13 and 14 show examples of control when the amount of reformed fuel (low ignitable fuel) generated by the fuel reformer 40 with a constant intake pressure is sufficient and when the amount is insufficient, respectively. It is a timing chart.

  When the amount of reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficient, the first combustion mode and the second combustion mode are switched at a predetermined switching load as shown in FIG. Perform switching.

  On the other hand, when the octane number (RON) in the reformed fuel (low ignitable fuel) generated by the fuel reformer 40 is sufficient but not sufficient, the low value as shown in FIG. Even if the ignitable fuel (high RON fuel) is increased, the average RON cannot be made sufficiently high, so this is dealt with by increasing the EGR rate. The EGR rate once increased stepwise is gradually decreased as the load increases.

  By increasing the RGR rate, the ignition delay is increased, the center of heat generation is retarded, and the sound vibration is reduced. In addition, since the thermodynamic MBT (optimum ignition timing) cannot be taken, the illustrated efficiency is lowered.

  In the above-described embodiment, the internal combustion engine 1 includes the fuel reformer 40. However, for example, if the reformed fuel tank 16 is configured so that gasoline can be supplied from the outside, the above-described fuel can be obtained. The reformer 40, the aggregator 45, and the like can be omitted.

  The technical ideas of the present invention that can be grasped from the above-described embodiments will be listed together with their effects.

  (1) A control device for an internal combustion engine has a fuel supply means that can directly supply a plurality of fuels having different ignitability at an arbitrary ratio according to a load to a combustion chamber, and compresses an air-fuel mixture that is leaner than a theoretical mixture ratio A first operation mode in which a low ignitable fuel and a high ignitable fuel are mixed in a combustion chamber in an internal combustion engine that operates by self-ignition combustion and has a control in which the ratio of fuel with low ignitability increases as the load increases. And a second operation mode in which the low ignitable fuel and the high ignitable fuel are distributed in different spaces in the combustion chamber, and the load when the amount of the low ignitable fuel exceeds a predetermined value is defined as a boundary. In addition, the first operation mode and the second operation mode are switched. In the second operation mode, there is a fuel amount necessary for the low ignitable fuel to be able to form a relatively homogeneous mixture with λ = 2 to 3, and the second amount of fuel is less than the load at which the fuel amount is reached. When the operation mode is performed, the air-fuel mixture becomes excessively diluted (λ> 3), and the unburned ratio increases, so that the thermal efficiency deteriorates. Moreover, even if the load for switching to the second operation mode is too high, it becomes difficult to suppress combustion noise. As a result, an optimal switching time between the first operation mode and the second operation mode is set, so that deterioration of thermal efficiency can be prevented and combustion noise can be suppressed.

  (2) In the control device for an internal combustion engine according to (1), the predetermined value is set to increase as the combustion chamber volume at the top dead center increases. In the second operation mode, there is a fuel amount necessary for the low ignitable fuel to be able to form a relatively homogeneous mixture with λ = 2 to 3, and the load that becomes the fuel amount depends on the shape of the combustion chamber. It is determined.

  (3) The control apparatus for an internal combustion engine according to (1) or (2), further including a variable compression ratio mechanism capable of changing a compression ratio of the internal combustion engine to an arbitrary compression ratio, wherein the predetermined value is a compression The ratio is set so as to decrease as the ratio increases.

  (4) The control apparatus for an internal combustion engine according to (1) or (2) described above includes a variable compression ratio mechanism that can change the compression ratio of the internal combustion engine to an arbitrary compression ratio. The load for switching between the first operation mode and the second operation mode is increased. Although it is desired to suppress the combustion noise by the second operation mode, the combustion noise is suppressed by lowering the compression ratio under the condition that the switching load between the first operation mode and the second operation mode has not been reached.

  (5) In the control device for an internal combustion engine according to any one of (1) to (4), the piston that defines the combustion chamber has a cavity in a piston crown surface, and the predetermined value is The cavity is set so as to increase as the volume of the cavity increases. Since the cavity is formed in the piston crown surface, it becomes easy to distribute the low ignitable fuel and the highly ignitable fuel separately in different spaces in the combustion chamber.

  (6) In the control device for an internal combustion engine according to any one of (1) to (4), the piston that defines the combustion chamber has a cavity in a piston crown surface, and the predetermined value is The smaller the volume outside the cavity of the combustion chamber at the top dead center, the smaller it becomes.

  (7) The control apparatus for an internal combustion engine according to any one of (1) to (6), further including fuel reforming means capable of controlling fuel properties, wherein the predetermined value is generated by the fuel reforming means. It is corrected according to the amount of stored fuel.

  (8) In the control device for an internal combustion engine according to (7), the load of switching between the first operation mode and the second operation mode increases as the amount of reformed fuel stored decreases.

  (9) The control apparatus for an internal combustion engine according to any one of (1) to (8), further including an intake pressure changing unit capable of arbitrarily controlling an intake pressure, wherein the amount of low ignitable fuel is the predetermined value. The intake pressure is fixed or lowered until the load is exceeded. Although it is desired to suppress the combustion noise by the second operation mode, the combustion noise is suppressed by lowering or fixing the intake pressure under the condition where the switching load is not reached.

8 ... Combustion chamber 9 ... Piston 10 ... Piston crown 11 ... Cavity

Claims (9)

It has a fuel supply means that can supply multiple fuels with different ignitability directly to the combustion chamber at an arbitrary ratio according to the load. In an internal combustion engine having a control in which the ratio of fuel with low ignitability increases with
A first operation mode in which low-ignition fuel and high-ignition fuel are mixed in the combustion chamber; a second operation mode in which low-ignition fuel and high-ignition fuel are distributed in different spaces in the combustion chamber; Have
A control apparatus for an internal combustion engine, wherein the first operation mode and the second operation mode are switched at a load when the amount of low ignitable fuel exceeds a predetermined value.   2. The control apparatus for an internal combustion engine according to claim 1, wherein the predetermined value is set to increase as the combustion chamber volume at the top dead center increases.   3. The variable compression ratio mechanism capable of changing the compression ratio of the internal combustion engine to an arbitrary compression ratio, wherein the predetermined value is set so as to decrease as the compression ratio increases. Control device for internal combustion engine.   It has a variable compression ratio mechanism that can change the compression ratio of the internal combustion engine to an arbitrary compression ratio, and the load for switching between the first operation mode and the second operation mode increases as the compression ratio decreases. The control device for an internal combustion engine according to claim 1, wherein   The piston that defines the combustion chamber has a cavity in a piston crown surface, and the predetermined value is set so as to increase as the volume of the cavity increases. The control apparatus of the internal combustion engine in any one.   The piston that defines the combustion chamber has a cavity in the piston crown surface, and the predetermined value decreases as the volume outside the cavity of the combustion chamber at the top dead center increases. The control apparatus for an internal combustion engine according to any one of claims 1 to 4.   7. A fuel reforming unit capable of controlling fuel properties, wherein the predetermined value is corrected according to a storage amount of fuel generated by the fuel reforming unit. A control device for an internal combustion engine according to claim 1.   The control device for an internal combustion engine according to claim 7, wherein a load for switching between the first operation mode and the second operation mode increases as the amount of reformed fuel stored decreases.   An intake pressure changing means capable of arbitrarily controlling the intake pressure is provided, and the intake pressure is fixed or lowered until a load is reached when the amount of low ignitable fuel exceeds the predetermined value. The control apparatus of the internal combustion engine in any one of 1-8.

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